Hydraulic control circuit for a variator

ABSTRACT

A variator assembly comprising a variator of the toroidal-race rolling-traction type. The assembly has input and output discs and rollers which are acted on by hydraulic roller control actuators. They are positioned between the discs to transmit torque from one disc to the other. Hydraulic end loading means supplied with fluid at an end load pressure to apply an end load to bias the discs and the rollers toward each other, thereby enabling the transmission of torque. Reaction pressure supply means are connected to the roller control actuators to cause them to apply an adjustable reaction force to the rollers. The assembly also has hydraulically influenced valve means which serve to compare an input related to the end load pressure with an input related to the reaction pressure and to control the end load pressure in dependence upon the comparison, thereby maintaining a relationship between the end load and reaction pressures. Adjustment means are provided to adjust the relationship between the end load and reaction pressures.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a hydraulic circuit for controlling acontinuously variable ratio unit (“variator”) of the toroidal-racerolling-traction type, and more particularly for controlling an end loadin such a variator.

2. Background Art

Toroidal-race rolling-traction type variators are in themselves wellknown. One, or more typically two, toroidal or part toroidal cavitiesare defined by opposed faces of rotatably, coaxially mounted discs anddrive is transmitted between the discs by rollers disposed in thecavities. It is well known in such variators to mount each roller in acarriage and to connect that carriage to a piston subject to acontrolled hydraulic force. So-called “torque control” operation can beachieved in well known manner by applying the hydraulic force along agenerally tangential direction (with respect to the axis of the variatordiscs) and allowing the roller/carriage to move along a circular pathcentred on the axis. The roller is permitted to precess (that is, theroller axis can rotate) and as is well known the roller precesses suchthat its axis always intersects the disc axis. Consequently when theroller moves along its circular path it also precesses and the change inroller inclination produces a change in variator transmission ratio. Theroller adopts a position in which the force applied thereto by thepiston is balanced by an opposite, “reaction”, force produced (by shearof a film of so-called “traction fluid”) at the interfaces between theroller and its neighbouring discs. The torque transmitted by thevariator is a function of the reaction force. In the steady state thehydraulic and reaction forces balance.

In order to enable transmission of torque by the variator there must bepressure at the roller/disc interfaces and in variators of the “fulltoroidal” type this is typically provided by means of a hydraulicactuator which acts on one of the variator discs to apply an “end load”biasing the discs toward the rollers. The magnitude of the end load hasan important bearing on variator efficiency and performance. It is knownto vary the end load during operation. An important parameter in thisregard is the traction coefficient. If we define the normal force to bethe force exerted by the roller on one of the discs (and of course bythe disc on the roller) at the interface therebetween and along thedirection normal to this interface, then the traction coefficient μ issimply the ratio of reaction force (RF) to normal force (NF):

$\mu = \frac{RF}{NF}$

Note that the normal force is in the general case not precisely equal tothe end load because the normal force acts along a directionperpendicular to the plane of the roller/disc interface, this directionbeing parallel to the direction of action of the end load only in oneparticular roller position (corresponding to a 1:1 variator driveratio). In the general case, the end load and normal force are relatedthrough the cosine of the roller angle.

An excessively low traction coefficient, corresponding to anunnecessarily high end load and hence high normal force, gives rise tolarge energy losses at the roller/disc interface and so is inefficient.An excessively high traction coefficient is also inefficient in energyterms and can lead to variator failure, excessive slip at theroller/disc interface allowing the roller to move, rapidly in somesituations, away from its proper position. It is necessary to guardagainst this eventuality.

In so called “full toroidal” variators, energy losses at the roller/discinterface can be considered in terms of (1) slip and (2) spin. Slipinvolves relative motion, along the circumferential direction, of theroller/disc surfaces at their interface, corresponding to a mismatch inrotational speeds of the roller and disc. Slip losses increase as thedegree of slip increases. Spin involves relative angular motion of thetwo surfaces at the roller/disc interface. It arises from the geometryof the variator and the degree of spin is determined by this geometry,the roller positions and the variator speed. However energy losses dueto spin are affected by the magnitude of the normal force and hence arerelated to the traction coefficient. It is found that the curverepresenting variation of efficiency with traction coefficient has apeak representing the best compromise between spin and slip losses. Thismust be taken into account in order to operate the variator at optimalefficiency.

A known hydraulic circuit for controlling the variator uses a pair ofhydraulic lines to supply hydraulic fluid at adjustable pressures toopposite sides of the roller control pistons, thus enabling the reactionforce to be varied. In order to provide for adjustment of end load, avalve arrangement of “higher pressure wins” type is used to supply fluidfrom whichever of the lines is at higher pressure to a working chamberof a hydraulic end load actuator and in this way a relationship iscreated between the reaction force and the normal force (or, to bestrictly accurate in view of the cosine variation of normal force withroller angle referred to above, between the reaction force and the endload). One such arrangement is described in the applicant's earlierEuropean Patent EPO894210 and in its US counterpart U.S. Pat. No.6,030,310 which disclosed in detail a practical end loading arrangementand the contents of which are incorporated herein by reference forpurposes of US law. In that arrangement the end load actuator actuallyhas two working chambers, one supplied with pressure from the higherpressure line to apply the end load and one supplied from the lowerpressure line to produce an opposed force which reduces the end load. Insuch an arrangement the traction coefficient can in effect be pre-set byappropriate choice of piston areas, particularly in the end loadactuator.

The hydraulic coupling of the end load to the roller control actuatorsmakes it possible to vary the end load rapidly in sympathy with thereaction force. This hydraulic coupling is highly advantageous becausevariators in motor vehicle transmissions are subject in practice torapid and severe “torque spikes”, eg. upon braking, and to provideadequate end load on demand to accommodate such spikes (and avoidvariator failure due to the traction coefficient increasing excessively)requires correspondingly rapid end load adjustment. In the arrangementdescribed above, occurrence of a torque spike results in a correspondingpressure increase in the higher pressure line which is automatically andrapidly passed on to the end load actuator by the hydraulics.

However such systems are subject to problems. In some arrangements poorpressure response, in particular a time lag in matching the variator endload to the roller reaction force, has been found to occur. Unavoidably,compliance in the variator and its hydraulics mean that a finite volumeof fluid is required to effect a change in end load. Flow is absorbed,eg. by flexure of the end load actuator components. In conjunction withflow restrictions in the hydraulics, the result can be a significanttime lag between the reaction pressure and the end load pressure andhence a transient mismatch between the end load and reaction forces. Themismatch occurs during rapid changes in reaction force as in the eventof transient torque spikes. In extreme cases there is an associated riskof variator failure.

It should be noted that EP0894210 suggests an arrangement in which ahydraulically controlled valve is used to control the end load pressure.This valve has a spool which is influenced by the end load pressureitself and also by mutually opposed pressures from opposite sides of thevariator's double acting roller control pistons. The spool's position isdetermined by the balance between these three signals. The end loadactuator is normally connected to a pump supplying pressurised fluid andthe valve controls a drain from the end load actuator so that inresponse to excess end load pressure the drain is opened and thepressure is reduced. The arrangement is intended to maintain thetraction coefficient at a constant level and there is no provision foradjustment of the traction coefficient.

It is desirable to provide for controlled adjustment of the tractioncoefficient, to make possible increased efficiency and to take accountof variable factors such as the temperature of the variator tractionfluid. Following start up the traction fluid, initially cold, isprogressively warmed and its characteristics are consequently altered.The appropriate traction coefficient is likewise altered and it would beadvantageous to carry out corresponding modification of the end load.

This need to adjust the traction coefficient according to temperaturehas been recognised in prior U.S. Pat. No. 6,162,144, assigned toGeneral Motors Corporation, although the hydraulic circuit used toachieve the adjustment (see FIG. 3 of the patent) simply uses a pulsewidth modulated valve to feed a percentage of the end load pressure to asecond chamber of the end load actuator, working in opposition to theend load pressure, to thereby adjustably reduce the end load. Theadditional problem of time lag in adjustment of the end load is notaddressed. Additionally it is believed that there would be severedifficulties in creating a practical implementation of the circuitproposed in this patent, particularly in providing a pulse widthmodulated valve capable of carrying out the required function.

It should also be noted that adjustment of the coefficient of tractioncan be achieved in the type of known hydraulic circuit discussed above,having two hydraulic supply lines feeding opposite sides of the rollercontrol pistons and a higher pressure wins arrangement to feed pressurefrom one of the lines to the end load actuator, by adjusting thepressures in both lines together to thereby increase or decrease thehigher pressure (and hence the end load) without altering the pressuredifference between the two lines which determines the reaction force.However this approach does not address the problem of end load time lagand complicates the control of the variator rollers.

SUMMARY OF THE INVENTION

The inventors have recognised that to address the two problems of endload time lag and traction coefficient adjustment requires a dual modeof end load pressure control, not found in the above mentioned priorart.

In accordance with a first aspect of the present invention there is avariator assembly comprising a variator of the toroidal-racerolling-traction type having input and output discs, rollers which areacted on by hydraulic roller control actuators and are positionedbetween the discs to transmit torque from one disc to the other,hydraulic end loading means supplied with fluid at an end load pressureto apply an end load to bias the discs and the rollers toward each otherthereby enabling the transmission of torque, and reaction pressuresupply means connected to the roller control actuators to cause them toapply an adjustable reaction force to the rollers, the assembly furthercomprising hydraulically influenced valve means responsive to thereaction pressure and the end load pressure to control the end loadpressure and thereby maintain a relationship between the end load andreaction pressures, and adjustment means to adjust the relationshipbetween the end load and reaction pressures.

In a preferred embodiment the effect of the adjustment means is toreduce the end load pressure so that in the event of inaction of theadjustment means the end load pressure is increased.

Preferably, the valve means serves to compare an input related to theend load pressure with an input related to the reaction pressure and tocontrol the end load pressure in dependence upon the comparison.

In a further preferred embodiment of the present invention the valvemeans comprise a pilot operated valve receiving a hydraulic reactionpressure input signal taken from a connection to the reaction pressuresupply means.

In another preferred embodiment of the present invention the rollercontrol actuators are double acting and are connected to first andsecond reaction pressure supply means, the pressures from which opposeeach other in determining the force applied to the rollers, a furthervalve arrangement being connected across the first and second supplymeans to direct the higher of the two pressures to the valve means.

Preferably the valve means is arranged to receive as a further input anend load adjustment signal from the adjustment means and to modify theend load pressure in response thereto.

In one such embodiment the valve means comprises a valve spool and theadjustment means comprises an actuator for applying an adjustablebiasing force to the valve spool.

The actuator may be coupled to the valve spool through a spring member.

In a further preferred embodiment the assembly further comprises apressure modifying arrangement which receives as an input pressure oneof the reaction pressure and the end load pressure, which modifies thispressure to create an output pressure which is a function of the inputpressure and which applies the output pressure to the valve means.

Preferably the pressure modifying arrangement comprises two restrictorswhich are connected in series and through which the input pressure isled to a pressure sink, one of the restrictors being variable and theoutput pressure being taken from between the two restrictors.

The valve means may have at least two states in which it serves toconnect the end loading means respectively to

(1) a high pressure fluid source and

(2) a pressure sink.

Preferably the valve means has a further state in which it serves toisolate the end loading means.

It is especially preferred that the adjustment means are electronicallycontrolled in dependence upon measured variator operating parameters.

According to a second aspect of the present invention there is avariator assembly comprising a variator of the toroidal-racerolling-traction type having input and output discs, hydraulic pistonactuated rollers positioned between said discs and operative to transmittraction therebetween and end loading means for applying hydraulicpressure to bias the discs towards engagement with each other, theassembly further comprising:

(a) means for supplying a reaction circuit pressure which is a functionof pressure applied to the roller pistons;

(b) an accumulator containing hydraulic fluid at an accumulatorpressure;

(c) pilot valve means switchable between a first position in which itconnects the reaction circuit pressure to the end loading means and asecond position in which it connects the accumulator to the end loadingmeans, the valve means being biased towards the first position by aforce which is a function of the pressure in the end loading means andbeing biased towards the second position by a force which is a functionof the reaction circuit pressure.

According to a third aspect of the present invention there is a variatorassembly comprising a variator of the toroidal-race rolling-tractiontype having input and output discs, hydraulic piston actuated rollerspositioned between said discs and operative to transmit tractiontherebetween and end loading means for applying hydraulic pressure tobias the discs towards engagement with the rollers, the assembly furthercomprising:

(a) means for supplying a reaction circuit pressure which relates topressure applied to the roller pistons;

(b) a pressurised fluid source; and

(c) pilot valve means switchable between a first position in which itconnects the reaction circuit pressure to the end loading means and asecond position in which it connects the pressurised fluid source to theend loading means, the valve means being biased towards the firstposition by a force which is related to the pressure in the end loadingmeans and being biased towards the second position by a force which isrelated to the reaction circuit pressure.

BRIEF DESCRIPTION OF THE DRAWINGS

Specific embodiments of the present invention will now be described, byway of example only, with reference to the accompanying drawings, inwhich:

FIG. 1 is a cross sectional view of a variator incorporating arepresentative, somewhat simplified, form of end load mechanism,together with a schematic representation of a control circuit for thevariator, incorporating certain features of the present invention;

FIGS. 2 to 7 are schematic representations of further control circuitsembodying certain features of the present invention;

FIGS. 8, 9 and 10 are schematic representations of still further controlcircuits, which embody the present invention; and

FIG. 11 is a graph illustrating operation of the invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

A variator 2 schematically illustrated in FIG. 1 comprises a pair ofinput rotor discs 78,79, an output rotor disc 4 and a plurality ofrollers 12 situated therebetween for transmission of torque in a mannerwell known to those skilled in the art and therefore not described indetail herein. At the end of the variator 2 there is provided anend-load assembly 5 which, in its representative form, comprises asimple hydraulic chamber 6 fed with hydraulic fluid at pressure. Thepressure in chamber 6 acts to load disc 78 axially such that it clampsthe rollers 12 between the discs 4, 78, 79 and enables transmission oftorque across the variator. As mentioned above, the magnitude of the endload is to be adjusted in order to achieve an appropriate tractioncoefficient.

Turning now to the control circuit 7 of the present embodiment of theinvention, it will be appreciated that the axle 10 of a master roller12′ of the variator is mounted in the cavity 14 of a hollow shaft 16 ofa double-headed roller carriage piston 18. This piston is formed withopposed piston heads 20, 21 which slide under hydraulic load withincoaxial cylindrical caps 23, 24 and which are free to rotate about theaxis of shaft 16. In practice a double-acting piston is oftenpreferable, with opposite faces of the single head both exposed tofluid, but the equivalent double-headed piston arrangement representedin the drawing has been used for ease of understanding. In either case,the piston reaction force is dependent on the difference in thehydraulic pressures applied to the two faces of the piston.

Hydraulic fluid inlets 26, 27 and outlets 29, 30 are formed in the endand side walls of caps 23, 24 respectively, and the end caps of a set ofslave roller carriage pistons 18 are supplied with fluid by means of aplurality of similar supply branches 25, 25 a via restrictors 31, 31 ain the supply branches 25, 25 a. The pressures acting on thecorresponding slave carriage pistons 18 of the remaining rollers arerelated to those in caps 23, 24 so that at equilibrium the appliedreaction forces equate.

The control circuit comprises two sources of hydraulic fluid provided byoil pumps 32, 33 capable of delivering hydraulic fluid from a sump 35at, for example, between 0 to 50 bar to left-hand and right-handupstream flow lines 37 and 38 and it is these lines that deliver thefluid respectively to the cylinder inlets 26 and 27 of the master piston18′ and to the slave pistons 18. Such pumps will, however, not providehydraulic fluid at these pressures unless control valves 58, 59,connected to the respective hydraulic outlets 29,30, are sufficientlyrestricted. A cross-connection 43 between lines 37 and 38 communicatesby way of a “higher-pressure-wins” arrangement of non-return valves 45and 46, and via a conduit 48, with a further control circuit 100 whoseoutput is connected to the hydraulic chamber 6 of the end-load mechanism5. This ensures that the further control circuit 100 is always fed withfluid at a pressure whichever of the two pressure lines 37, 38 is ofhigher pressure (hereinafter referred to as “reaction circuitpressure”).

Outlets 29 and 30 from caps 23 and 24 lead by way of downstreamleft-hand and right-hand lines 55 and 56 to the inlets of the twopressure control valves 58 and 59 which are formed as electro-hydraulicproportional pressure control valves and whose operation is describedlater herein. Downstream of the control valves 58 and 59, the left andright-hand fluid lines combine at 68 after which a connection 70 isoperable to provide fluid for general lubrication of the transmission.This is maintained at the correct back pressure by pressure relief valve72.

The further control circuit 100 comprises a shuttle valve 102 in theform of a double pilot operated directional control valve having a firstinlet port for receipt of the reaction circuit pressure from the“higher-pressure-wins” valve arrangement 45, 46 via an end-loaddischarge dump valve 104, a second inlet port connected to a highpressure fluid source formed in this embodiment as a hydraulic pressureaccumulator 106, and an outlet port connected directly to the hydraulicchamber 6 of the end-load mechanism 5. The discharge dump valve 104 is aflow control valve which allows free flow in the forward direction (i.e.towards the control circuit 100) but which regulates and limits reverseflow back to the reaction circuit in that the valve dumps any excessreverse flow above a predetermined level (e.g. the valve allows amaximum reverse flow of typically 0.5 liters/min and dumps any excessflow).

The shuttle valve 102 is biased by a spring 108 towards the positionillustrated in FIG. 1, in which the output of the discharge dump valve104 is connected to the outlet of the shuttle valve 102 and thence tothe hydraulic chamber 6 of the end-load mechanism 5. However, theposition of the valve is further determined by two pilot pressures. Thefirst pilot pressure is taken by line 110 directly from the hydraulicchamber 6 of the end-load mechanism 5 and corresponds to the hydraulicpressure existing in the chamber 6 at any time. The second pilotpressure is taken from the higher of the two pressures applied to one ofthe slave pistons 18, by means of a zero flow shuttle valve 112 via line111. The valve 112 is actuated by fluid pressure at points between thepiston outlets and the flow restrictors 31, 31 a, preferably as close aspossible to the piston outlets, since these pressures are mostrepresentative of the actual pressures existing within the cylinder 18′.

As will be appreciated by those skilled in the art, the valve 112 may beactuated by pressures from opposite sides of two different slave pistons18, since the pressures applied to the same side of each of the pistons18 are the same. In the drawing, the pressures are taken from the twoends of the same piston, for ease of reference. Alternatively, the valve112 may be actuated by pressures from opposite sides of the masterpiston 18′, but this may cause problems if the piston is fitted with ahydraulic end stop mechanism.

The two pilot pressures act in opposite senses on the pilot-operatedshuttle valve 102 and thus the shuttle valve 102 serves as a comparator,comparing the higher of the pressures applied to the variator controlcylinders with the end-load pressure. Since the shuttle valve 102 reactsto pressures within one of the slave pistons 18 as restricted by theflow restrictors 31, 31 a and since the pilot signal from the variatorcontrol cylinder requires no flow, the output from the valve 112 istherefore the most accurate indicator of pressure, and thereforereaction force, within the cylinders. In particular, it overcomes thepotential of a false indication of pressure resulting from pressurelosses (and therefore time lags) which occur when the end-load is beingcharged and slower pressure rise in the hydraulic chamber 6 of theend-load mechanism 5.

As mentioned above, the shuttle valve 102 has a bias provided by spring108, such that the end-load pressure must fall below the maximumreaction circuit pressure by an amount in excess of the preloading ofthe spring 108 (typically in the region of 1 bar) before the shuttlevalve moves to its alternative position in which the accumulator 106 isconnected to the end-load.

Thus, for situations where the end-load pressure exceeds the reactioncircuit pressure, and also where the reaction circuit pressure exceedsthe end-load pressure by less than the bias of the spring 108, theshuttle valve will connect the output of the “higher pressure wins”valve arrangement 45, 46 to the chamber 6 of the end-load mechanism 5.This ensures accurate steady state pressure matching of the reactioncircuit pressure from the variator and the end-load pressure and makesthe system determinate under steady state conditions.

When the end-load pressure falls below the reaction circuit pressure byan amount in excess of the preloading of the spring 108, thepilot-operated shuttle valve 102 moves to its alternative position,connecting the chamber 6 of the end-load assembly 5 to the accumulator106 (which is typically charged to a pressure of about 50 to 55 bar).The end-load chamber 6 is then charged from the accumulator, but becauseof the differential area of the spool of the valve 102 at full strokethe end-load must rise above reaction circuit pressure before the valve102 moves back to connect the end-load chamber 6 to the reaction circuitfeed.

In practice, it is found that the shuttle valve 102 tends to shuttleback and forth between its two extreme positions in transientconditions. The valve rarely remains connected to the accumulator for asufficient length of time for the accumulator pressure to be appliedfully to the end load chamber 6, since whenever a different pressure isapplied to the end load chamber 6, the pressures controlling theposition of the shuttle valve are also changed as a consequence.

The result of the above is that during a transient situation theend-load is applied stepwise as the pilot-operated shuttle valve 102shuttles back and forth to top up end-load pressure as required.Reaction to transient “spikes” in the reaction pressure is rapid. Thesystem allows the use of small mechanical valves and is therefore verydynamic.

If pressure demand in the end-load assembly drops (as a result of a fallin the pressures applied to the variator control pistons), any excess inthe flow (over a predetermined limit) from the end-load chamber 6 isdumped to tank via the end-load dump valve 104 rather than flowing backinto the reaction circuit, thereby ensuring good pressure-off times.

FIGS. 2 to 10 illustrate various further variator control circuits in amore schematic form, only those components directly related to end loadcontrol being included. The same reference numerals are used throughoutFIGS. 2 to 10 to indicate certain common components. Refer firstly toFIG. 2 which includes all such numerals. The end load actuator is seenat 200 and as before has a working chamber 202 supplied with fluid at anoperating pressure through a shuttle valve 204 formed as a double pilotoperated directional control valve which in each case receives ahydraulic pilot signal representative of the end load operatingpressure. In each of the embodiments illustrated in FIGS. 2 to 10 thisis provided through a passage 206 which is additionally connected to afirst port 207 of the shuttle valve and to the working chamber 202.FIGS. 2 to 10 all show only a single representative roller controlpiston 210, shown schematically in these illustrations within a cylinder212 to form a double acting arrangement, opposite faces of the piston210 receiving hydraulic fluid through respective branches 214, 216corresponding to the branches 25, 25 a of the FIG. 1 circuit. Thosecomponents of the hydraulic circuit used to generate adjustable pressurein the branches 214,216 in order to control the roller reaction forceare omitted from FIGS. 2 to 10, but may be as in FIG. 1. The branches214, 216 incorporate respective flow restrictions 218,220 whose functionis to provide damping of roller motion. A higher pressure wins valvearrangement 222 has respective inputs connected on either side of therepresentative roller control cylinder 212. These connections are formedbetween the cylinder and the respective restrictors 218,220, in orderthat the valve arrangement 222 can pass on, through a conduit 225, thehigher of the reaction circuit pressures with minimal time lag to serveas a second pilot pressure signal acting on the shuttle valve 204 andopposing the signal representing the end load pressure. In each of FIGS.2 to 10, and similarly to the FIG. 1 embodiment, a second port 223 ofthe shuttle valve 204 is connected to a high pressure fluid sourcecomprising a hydraulic accumulator 224 which is maintained at therequired pressure by means of a pump 226, a relief valve 228 and a nonreturn valve 230 in a manner with which the person skilled in the artwill be familiar.

The circuits illustrated in these drawings differ from each other, amongother things, with regard to the connection of the shuttle valve's thirdport 232. In each case however the shuttle valve 204 is a three positionvalve which serves, in its different positions, to:—

i. isolate each port from the other, as in the position shown in thedrawings;

ii. connect the first and second ports 207, 223 to deliver hydraulicfluid pressure from the accumulator 224 to the working chamber 202 ofthe end load actuator 200; or

iii. connect the first and third ports 207, 232 to connect the end loadworking chamber 202 to some part of the circuit serving as a pressuresink.

In all of these circuits the path to the pressure sink incorporates aflow restrictor 234 whose function will be described below.

Looking now specifically at FIG. 2, it can be seen that the third port232 of the shuttle valve 204 is connected through the flow restrictor234 to a drain leading to the sump 240. There is no provision forconnecting the reaction circuit pressure to the end load working chamber202, as in the FIG. 1 circuit. Nonetheless the reaction circuit pressuredoes have a controlling influence on the end load pressure since theshuttle valve 204 serves to compare the two pressures, due to its pilotinputs. If the reaction circuit pressure overcomes the end loadpressure, indicating that the end load is insufficient, then the valve204 shuttles to position (ii) referred to above (spool in its right-mostposition) to connect the accumulator 204 to the end load actuator 200and so increase the end load, until the required pressure is achieved atwhich point the valve shuttles back to position (i) to maintain end loadat a steady level. If on the other hand reaction circuit pressure fallsto a sufficiently low level relative to the end load pressure then thevalve shuttles to position (iii) allowing discharge of the end loadworking chamber 202 through the flow restrictor 234, until once more theend load and reaction circuit pressures acting on the shuttle valve 204balance such as to allow it to return to position (i).

The flow restrictor 234 controls the end load decay rate, ensuring goodend load fraction during repeated transients and imposing a limit on theflow required from the accumulator 224 no matter how high the frequencyof disturbance to the system. Above a certain frequency of excitation,the end load simply remains high. In between the steady state conditionand the infinite frequency case lies a maximum mean accumulator flow,occurring at a certain frequency and magnitude of pressure disturbance.The capacity of the high pressure fluid source 224, 226 can be selectedon this basis.

Note that the comparison between the end load and reaction circuitpressures, effected by the shuttle valve 204 in this exemplaryembodiment, can be weighted. The areas of the valve's spool subject tothe two pilot pressures need not be equal, and the ratio of one to theother can be used to set the traction coefficient. Also the valve'sspool is typically biased, eg. by mechanical springing.

The circuit illustrated in FIG. 3 differs from that of FIG. 2 in thatthe pressure sink is provided by means of a conduit 250 leading from theflow restrictor 234 to the higher pressure wins valve arrangement 222and hence to the higher pressure side of the roller control circuit.

The circuit illustrated in FIG. 4 corresponds to the FIG. 3 circuitexcept that a flow control (regulating) valve 260 is used in the routeto the pressure sink (again, provided through conduit 250 leading to theroller control circuit through the valve arrangement 222) in place ofthe flow restrictor 234. The flow control valve 260 controls the rate ofend load discharge and sends a regulated flow through the conduit 250 tothe roller control circuit, most of the end load discharge flowtypically being dumped by the valve 260 to the sump through a dumppassage 262. This prevents the reaction circuit pressure from beingdisrupted when the end load discharges.

The circuit illustrated in FIG. 5 differs from that of FIG. 3 in thatend load discharge is provided for by means of a further higher pressurewins valve arrangement 270 which receives the discharge fluid through aconduit 272 leading from the flow restrictor 234 and delivers it,through whichever of its outlets 274, 276 is at higher pressure, to themain roller control circuit, outboard of the relevant roller flowrestrictor 218, 220.

FIG. 6 illustrates a development of the circuit illustrated in FIG. 5,the difference between the two circuits being that the flow restrictor234 of FIG. 5 has been replaced by a flow control valve 280. Thisperforms a similar role to valve 260 of the FIG. 4 circuit, controllingthe rate of end load discharge and sending a regulated flow back to thereaction circuit during discharge, the remainder of the flow beingdumped to the tank through dump passage 282. This prevents disruption ofroller control circuit pressure upon end load discharge.

FIG. 7 illustrates a fail safe feature which may be added to any of theillustrated circuits. A back-up conduit 290 conducts the higher reactioncircuit pressure from the higher pressure wins arrangement 222 to anormally closed check valve 292, connected in its turn to the end loadactuator 200. If for any reason (eg. malfunction of the circuit) the endload should fall to an unacceptably low level relative to reactionpressure, which could otherwise create a danger of inadequate endloading leading to variator traction failure, then the check valve 292,is caused to open providing a low resistance path for the end load to becharged from the reaction circuit. An operational sensor 294, formed inthis embodiment as a differential pressure switch, would then indicateto the transmission's electronic control (PCU) that the problem hadoccurred, enabling appropriate control to be implemented to protect thetransmission, and in some embodiments providing the driver with awarning signal.

While the embodiments described above allow rapid and effective end loadcontrol, addressing the end load time lag problem which has already beendiscussed, it has yet to be explained how they may be adapted to permitadjustment of the relationship between reaction circuit pressure and endload, and hence adjustment of the variator traction coefficient. FIGS. 8and 9 illustrate embodiments in which provision is made to effectively“weight” the comparison of reaction circuit and end load pressures,thereby adjusting the relationship between the two pressures and hencethe traction coefficient which is achieved. In the illustratedembodiments, utilising the shuttle valve 204 to carry out thiscomparison, the weighting is effected by means of adjustable biasing ofthe shuttle valve spool, providing what is in effect a further controlinput thereto. This approach can be applied to any of the previouslydescribed circuits, as will be apparent from the following.

The circuit illustrated in FIG. 8 corresponds to that of FIG. 7 exceptthat a solenoid 300 acts on the spool of the shuttle valve 204 such asto apply an adjustable biasing force thereto, urging the spool towardthe end load discharge position (iii)—i.e. to the left, in the drawing.The magnitude of the biasing force is controlled by the transmission'sPCU 302. The shuttle valve 204 thus adopts the end load charge position(ii) only when the reaction circuit overcomes the sum of the end loadpilot force and the solenoid biasing force. This feature thus enablesthe PCU to introduce an adjustable negative offset on the end load levelin order to increase traction coefficient to an optimal level, therebypotentially increasing efficiency. This control can be carried out independence upon measured operating parameters such as operatingtemperature or indeed prevailing variator roller positions, received bythe PCU as inputs 301.

This mode of traction coefficient control may be referred to as“subtractive” since the additional, variable input to the shuttle valve204 provided by the solenoid 300 serves to reduce the end load. This hasthe benefit of being fail safe in the event of failure of the solenoid300. If the solenoid fails, applying no biasing force, the effect is toincrease the end load which reduces efficiency but still providesadequate end load for variator function, enabling the vehicle to “limphome”.

In FIG. 9 a solenoid 310 is provided which biases the shuttle valve 204in the direction (opposite to the direction of action of solenoid 300)toward the accumulator charge position (ii). Hence the valve onlyshuttles from the accumulator charge position when the end load pilotforce is sufficient to overcome the sum of the reaction pressure pilotforce and the biasing force applied by the solenoid 310. Again thesolenoid's biasing force is controlled by the PCU 302 which cantherefore set a positive offset on the end load level. This “additive”control can again be used to adjust the traction coefficient to provideefficiency improvements.

It will be apparent that any of the shuttle valves 204 of the circuitsillustrated in FIGS. 1 to 7 may receive an additional control signal,e.g. from a solenoid such as 300 or 310, to enable traction coefficientadjustment in accordance with the present invention.

The skilled reader will recognise that the function of the shuttle valve204 can be put into practice in a variety of ways. For example, it isdesirable not to directly secure the solenoid 302 to the valve spoolsince the solenoid is relatively massive and the valve's speed ofresponse would consequently be impaired. A more practical alternative isto couple the solenoid and spool through a spring, so that the solenoidserves to apply an adjustable mean force to the spool. A furtheralternative is to have the solenoid, or some other actuator, adjust theposition of a sleeve of the valve 204 defining the valve's ports inorder to provide the required offset.

The function of the solenoids 300, 310 is, as will be apparent, toadjust the relationship between the end load pressure and the reactioncircuit pressure. Such adjustment may however be achieved in other ways.One alternative is illustrated in FIG. 10. As in earlier circuits thepressure supplied to the working chamber 202 of the end load actuator200 is controlled by the three position shuttle valve 204 whose spool isinfluenced by a first pilot pressure signal from connection 206 to theworking chamber 202 and by an opposed, second pilot pressure signalderived from the higher pressure wins arrangement 222. However in theFIG. 10 circuit this second pilot pressure signal is adjustable. Insteadof being led directly to the spool of the shuttle valve 204, the outputof the higher pressure wins arrangement 222 is connected through aseries combination of first and second flow restrictors 350, 352 todrain. The second pilot pressure signal is taken through a conduit 354connected to a point between the first and second flow restrictors 350,352, one of which provides a variable restriction under control from thePCU 302. In FIG. 10 this function is carried out by the first flowrestrictor 350 which takes the form of an electronically controlledvalve.

The first and second flow restrictors function analogously to apotential divider in an electronic circuit. A pilot flow passescontinuously through the restrictors, this flow being small enough notto significantly alter the reaction pressure obtained from the higherpressure wins arrangement 222.

In flowing from the higher pressure wins arrangement to the tank (herelabelled 354, and of course being at atmospheric pressure) the fluidexperiences a total pressure drop equal to the reaction circuitpressure. Neglecting flow resistance in the intervening conduits, thispressure drop takes place across the two flow restrictors 350, 352. Theratio of the pressure drop ΔP₁ across first restrictor 350 to thepressure drop ΔP₂ across second restrictor 352 is determined by theresistance to flow of the two restrictors and (while the valve formingthe first restrictor 350 is not adjusted) is largely constant regardlessof changes in flow rate with variations in reaction circuit pressure.Adjusting the restrictor 350, however, allows the ratio of ΔP₁ to ΔP₂ tobe correspondingly adjusted.

Consequently an adjustable fraction of the reaction circuit pressure isapplied to the shuttle valve 204 to serve as the second pilot pressuresignal. Hence by controlling the variable restrictor 350, therelationship between the end load and the reaction circuit pressure canbe adjusted.

Whereas FIGS. 8 and 9 are concerned with “subtractive” and “additive”adjustment of this relationship, FIG. 10 shows a way to achieve what maybe referred to as “multiplicative” adjustment. FIG. 11 is a graphintended to make this distinction clear, reaction circuit pressure (orequivalently variator reaction torque) CP being shown on the horizontalaxis against end load EL on the vertical axis. The straight line Aindicates the (idealised) relationship between these two variablescreated by use of the shuttle valve 204 in the absence of adjustment bysolenoids 300, 310 or by variable restrictor 350.

The ratio of end load to reaction circuit pressure is constant. To putthis another way, neglecting the cosine factor referred to above, thestraight line A corresponds to a constant traction coefficient. Straightline B shows the effect of subtractive adjustment of end load aseffected by the FIG. 8 circuit. The gradient of the line relatingreaction and end load pressures is unchanged but the line is offsetvertically, no longer passing through the origin. Consequently there isno longer a constant ratio of one pressure to the other. The tractioncoefficient thus varies with changes in reaction circuit pressure. LineC shows an additive adjustment, again with a non-constant tractioncoefficient. Line D on the other hand shows a multiplicative adjustmentcarried out by the FIG. 10 circuit. The gradient differs from line A butthe line still passes through the origin, indicating that the effect ofthe adjustment is to change the traction coefficient to a new valuewhich is nonetheless constant notwithstanding changes in the reactioncircuit pressure.

The circuits illustrated in FIGS. 8, 9 and 10, it will be understood,utilise the hydraulically controlled shuttle valve to maintain arelationship between end load and reaction pressures. Beinghydraulically controlled the valve can have a quick response time and soreact with sufficient rapidity to maintain adequate traction even duringrapid “transients” such as in the event of rapid vehicle braking oracceleration, which briefly create very high variator torque demand. Thesolenoid 302 or the restrictor valve 350, serving to adjust thisrelationship, are electronically controlled and consequently slower inresponding but the required adjustment (corresponding e.g. to fluidtemperature) can be more slowly carried out without impairing variatorfunction.

1. A variator assembly comprising a toroidal-race rolling-tractionvariator having input and output discs, rollers which are acted on byhydraulic roller control actuators and are positioned between the discsto transmit torque from one disc to the other, a hydraulic end loadingdevice supplied with fluid at an end load pressure to apply an end loadto bias the discs and the rollers toward each other, thereby enablingthe transmission of torque, and a reaction pressure supply connected tothe roller control actuators to cause them to apply an adjustablereaction force to the rollers, the assembly further comprising ahydraulically influenced valve arrangement which serves to make acomparison of an input related to the end load pressure with an inputrelated to the reaction pressure and to control the end load pressure independence upon the comparison, thereby maintaining a relationshipbetween the end load and reaction pressures, and an adjustment devicewhich serves to adjust a signal applied to the valve arrangement andthereby to adjust the relationship between the end load and reactionpressures, the adjustment device being controlled by an electroniccontroller in response to measured operating parameters.
 2. The variatorassembly of claim 1 wherein an effect of the adjustment device is toreduce the end load pressure so that in the event of inaction of theadjustment device, the end load pressure is increased.
 3. The variatorassembly of claim 1 wherein the valve arrangement comprises a pilotoperated valve receiving a hydraulic reaction pressure input signaltaken from a connection to the reaction pressure supply.
 4. The variatorassembly of claim 1 wherein the roller control actuators are doubleacting and are connected to first and second reaction pressure supplies,the pressures from which oppose each other in determining the forceapplied to the rollers, a further valve arrangement being connectedacross the first and second supply means to direct the higher of the twopressures to the valve arrangement.
 5. The variator assembly of claim 1wherein the valve arrangement is arranged to receive as a further inputan end load adjustment signal from the adjustment device and to modifythe end load pressure in response thereto.
 6. The variator assembly ofclaim 5 wherein the valve arrangement comprises a valve spool and theadjustment device comprises an actuator for applying an adjustablebiasing force to the valve spool.
 7. The variator assembly of claim 6wherein the actuator is coupled to the valve spool though a springmember.
 8. The variator assembly of claim 1 further comprising apressure modifying arrangement which receives as an input pressure oneof the reaction pressure and the end load pressure, which modifies thispressure to create an output pressure which is a function of the inputpressure and which applies the output pressure to the valve arrangement.9. The variator of claim 8 wherein the pressure modifying arrangementcomprises two restrictors which are connected in series and though whichthe input pressure is led to a pressure sink, one of the restrictorsbeing variable and the output pressure being taken from between the tworestrictors.
 10. The variator assembly of claim 8 wherein the inputpressure to the modifying arrangement is the reaction pressure.
 11. Thevariator assembly of claim 3 wherein the valve arrangement has at leasttwo states in which it serves to connect the end loading devicerespectively to (1) a high pressure fluid source and (2) a pressuresink.
 12. The variator assembly of claim 11 wherein the valvearrangement has a further state in which it serves to isolate the endloading device.
 13. The variator assembly of claim 11 wherein thepressure sink comprises a drain at atmospheric pressure.
 14. Thevariator assembly of claim 11 wherein the pressure sink comprises aconnection to the reaction pressure supply means.
 15. The variatorassembly of claim 11 wherein the connection to the pressure sink isthrough a flow restrictor serving to limit throughput of fluid of thevalve means.